The performance, such as accuracy and repeatability, of a precision machine operating in a vibration environment can be improved if the vibration isolation system supporting the machine provides controllable stiffness and/or damping characteristics during the machine's various phases of operation.
As schematically shown in FIG. 1, a machine 5 has a machine base 29 which is supported, via a cradle 58, by a plural number of isolator assemblies 40 which rest on a support 48. A movable component 27 of the machine is driven by a motor 25 through a transmission mechanism 26 to move along direction 28. Motor 25 is controlled by a control signal 21 which is generated by a motion controller 20 upon receiving either a digital motion signal 15 from a digital computer 10 or an analog motion signal 14 from a manual controller 12. The motion of movable component 27 during a typical machine operation cycle serves two purposes: a) positioning movable component 27, and b) accomplishing an intended task. The velocity profile of movable component 27 during a typical operation cycle is depicted in FIG. 2(a) where the above mentioned first purpose is accomplished during phases A, B, and C, and the second during phase D. The speed of movable component 27 first increases at a uniform rate to V.sub.-- 2 during phase A and then stays constant during phase B as shown in FIG. 2(a). When movable component 27 is at a pre-determined distance from its destination, it's speed is reduced to V.sub.-- 1 during phase C. Movable component 27 then moves at a constant speed V.sub.-- 1, which can be zero for certain applications, during phase D till the intended task is accomplished.
High throughput demand for machine 5 requires that movable component 27 spend the least possible time during phases A, B, and C. This results in a large magnitudes for acceleration A.sub.-- 1 and, deceleration A.sub.-- 2 in phases A and C, respectively (FIG. 2(b)). A greater magnitude of acceleration and deceleration of movable component 27 in turn results in greater magnitude of forces, F.sub.-- 1 and F.sub.-- 2 in phase A and B, respectively. These forces are exerted on isolator assemblies 40 (FIG. 2(c)), causing machine base 29 to deviate from and oscillate about its equilibriums position, and causing a degraded machine performance during phase D. To minimize this deviation, it is preferable that isolator assembly 40 has a maximum stiffness characteristic during phases A, B, and C. It is also desirable for isolator assembly 40 to exhibit a maximum damping characteristic to facilitate settling of the oscillation before movable component 27 starts the intended task during phase D.
The high precision requirement for the machine's intended task at the end of phase D, however, requires maximum efficiency of vibration isolation of the machine by vibration of support 48. This requires that isolator assembly 40 exhibit minimum stiffness and damping characteristics, the former reduce the natural frequency of the system and the latter improve the isolation efficiency of the machine by vibration of support 48 at higher frequencies.
Clearly the high throughput and high precision demands for the machine impose conflicting requirements on the stiffness and damping characteristics of the vibration isolation system. However, as shall be discussed in detail below, no existing art can satisfy these requirements.
Most of the vibration isolation systems available today have fixed stiffness and damping characteristics. They are referred to as "semi-active" because they can automatically re-level a machine supported by them in response to a slow change in level of the machine due to a change in the force exerted on them. FIG. 1(b) shows schematically a cross-sectional view (along line 1b--1b in FIG. 1(a)) of such a system in which isolator assembly 40 is disposed, via cradle 58, between machine 5 and support 48. An adapter 41 connects cradle 58 and a flexible diaphragm 42 supported by pressurized gas in chamber 50. Chambers 50 and 49, both are charged by pressurized gas 45 via a gas passage 46, are separated by separator 51 and communicate with each other via a communication passage 54. Chamber 49 serves both as a reservoir and, together with communication passage 54, as a damper in isolator assembly 40. A pressure regulator 44 regulates the gas pressure in chambers 49 and 50, via a gas passage 47 and a communication passage 54, according to feedback from a lever 43 which responds to the deviation of cradle 58 from its equilibrium position caused by a change in force exerted on isolator assembly 40. The stiffness of isolator assembly 40 is determined by the size of chambers 50 and 49, while the damping is primarily determined by the gas flow rate through communication passage 54. Because of the low dynamic response of lever 43, pressure regulator 44, and gas in chambers 50 and 49, this mechanism only operates are required for slow changes in the exerted force, such as a change in force caused by a change in the center of gravity of the machine which results in a slow but significant change in the position of movable component 27. This type of vibration isolation system is inadequate for ensuring adequate performance of a precision machine because of its fixed stiffness and damping characteristics and its slow dynamic response. Additionally, components such as cradle 58 and its connectors introduce problems such as instability, inadequate connection between machine base 29 and isolator assembly 40, and design complexity.
A second type of vibration isolation system is referred to as "active" because it uses transducers, such as accelerometers, to constantly sense certain parameters of machine motion, such as velocity or acceleration, which are the result of external disturbances. The senses values of these parameters are used to control the generation of either a force or a displacement in response to the external disturbances for the purpose of maintaining the machine at its equilibrium position. This type of isolation system not only requires the use of sensors but also proved to be inadequate for ensuring adequate performance of a precision machine because of its limited capability for controlling its displacement, and stiffness. These capability limitations result in poor vibration isolation efficiency at high frequencies.
U.S. Pat. No. 4,757,980 issued to Dale W. Schubert discusses a parametrically controlled active vibration isolation system whose damping is controlled by using a servo valve which is actuated in response to velocities, measured by two velocity sensors. The sensors measure velocities of a machine and a support connected by the vibration isolation system. The system is intended for isolating the machine from vibration of the support. Similarly to all other prior art active isolation systems, it requires the use of sensors. Additionally, it is inadequate for application to a precision machine because it relies on velocity signals. As shall be discussed later, mechanical shocks and change in forces acting on isolator assemblies are primarily caused by acceleration, not velocity, of the movable component of a machine.
U.S. Pat. No. 5,379,980 issued to Worthington B. Houghton et al. discusses systems for intermittent stabilization of a machine by determining a load shift condition, increasing passage of gas to pneumatic isolators supporting the machine, and imposing external or internal damping to the machine in response to the load shift condition. These systems rely on sensors to determine or a computer to anticipate a load shift condition such as the position change of the payload or its loading or unloading to and from the machine. The drawbacks of these systems, in addition to having to use sensors or signals from a computer, include slow response and overshooting due to the slow dynamic response of gas, which fills the chambers of the pneumatic isolators, to the load shift condition. A somewhat similar system is disclosed in U.S. Pat. No. 4,941,265 issued to Klaus-Peter Heiland, wherein the position of a movable element of a machine, which is either monitored by sensors or derived from a digital computer, is compared with a memory for actuating a plunger to change the gas pressure in an isolator to maintain the position of the machine. It also includes logic means to ensure that the plunger is already actuated before anticipated motion of the movable element takes place. This system, in addition to the need for a complex circuitry to generate a driving signal for the plunger and to detect the position of the plunger, reduces the machine's throughput because of the slow dynamic response of gas being pressurized or released by the plunger. Therefore this kind of system can not be applied to counter a change in machine position due to dynamic forces such as those caused by the acceleration or deceleration of the movable element of the machine. In addition, the occurring of the plunger motion before the anticipated motion of the movable element tilts the machine in the opposite direction to that caused by the anticipated motion of the movable element, negating the purpose of maintaining the position of the machine. Besides, the position signal from the computer of a machine is not always executed immediately by the motion controller. Therefore the position signal, or any signal from the computer, is generally not a reliable indication of the motion status of the movable component of the machine.
There also exists prior are relating to the design of vibration dampers with controllable damping characteristic. Two kinds of dampers with certain relevance to the present invention are described in U.S. Pat. No. 5,277,281 issued to J. David Carlson et al., U.S. Pat. No. 5,573,088 issued to John J. Daniels. The U.S. Pat. No. 5,277,281 patent disclosed a controllable damper whose damping characteristic is varied by changing the strength of a magnetic field applied to a magnetorheological fluid, while the U.S. Pat. No. 5,573,088 patent accomplishes the same purpose by varying the strength of an electric field applied to an electrorheological fluid. Because of the limited load carrying capabilities and inadequacy in the amount of allowable deflection along multiple directions, improvements are needed before the above mentioned controllable dampers can be applied to precision machines operating in a vibration environment.